Mixed flow pump

ABSTRACT

A highly efficient mixed flow pump can prevent flow separation which is likely to occur in a corner portion of a flow passage of a diffuser section. The mixed flow pump includes a casing having an axis and defining an impeller section and a diffuser section disposed downstream of the impeller section with stationary diffuser blades protruding from a hub. The diffuser blades are formed so that an angular difference, between a hub blade angle and a casing blade angle, is chosen to conform to a specific distribution pattern along a flow passage of the diffuser section.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates, in general, to a mixed flow pump having adiffuser section with diffuser blades for guiding flow therein.

2. Description of the Related Art

A conventional mixed flow pump, shown in a cross sectional view in FIG.12, is comprised of a casing 16 housing an impeller 12 rotating about anaxis of a rotation shaft 10, and a stationary diffuser section 14disposed downstream of the impeller 12. The flow passage P in thediffuser section 14 is formed as a three-dimensionally curved space in aring-shaped space formed between the casing 16 and a hub 18, separatedby diffuser blades 20. A fluid medium taken through a pump inlet 22 isgiven kinetic energy by the rotating impeller 12, and is reduced of itscircumferential velocity as the fluid enters into the stationarydiffuser section 14, and the kinetic energy at the impeller exit isrecovered as a static pressure in the pumping system.

The shape of the flow passage P in the diffuser section 14 is definedaccording to the shape of the meridional (axisymmetrical) surfaces ofthe hub 18 and the casing 16 and the geometrical shape of the diffuserblades 20. Of these three, the shape of the blades is determined bychoosing a distribution pattern of blade angle β which is an anglebetween a direction M tangential to a center line of the blade on theaxisymmetrical surface of the hub 18 or the casing 16 at any given pointalong the blade length and the tangent L in the circumferentialdirection at that point, as illustrated in FIG. 13A.

The blade angle β is given by an equation relating the meridionaldistance m (defined by the distance along the line of intersection of aplane containing the rotation axis of the impeller 12 and theaxisymmetrical surface) and a circumferential coordinate θ and a radialcoordinate r for the blade center line as follows (refer to FIG. 13C):

 tan β=dm/d(rθ)  (1)

The blade angle β of the diffuser blade 20 at the entrance-side of thediffuser section 14 is chosen to coincide with the direction of thestream flow at the exit of the impeller 12, and the blade angle β of thediffuser blade 20 at the exit-side of the diffuser section 14 is chosenso that the exiting flow is produced primarily in the axial directionafter being eliminated of the circumferential velocity component of theflow. In the flow passage that lies between the entry and exit regionsof the diffuser section 14, it is a general practice in the conventionaldesign technology to adopt a smooth transition of blade angles resultingin that, as shown in FIG. 14A, the blade angle distribution pattern issimilar along the hub surface and along the casing surface. In theillustration shown in FIG. 14A, the non-dimensional distance m* isdefined by normalizing the meridional distance m by the distance l fromthe leading edge to the trailing edge of a blade along either the hubsurface or, the casing surface. FIG. 15 shows the blade angledistribution pattern of the blade angle difference Δβ between the hubblade angle and the casing blade angle in a conventional diffusersection operating in a specific speed range between 280˜700 (m, m³/min,rpm) with respect to the non-dimensional distance m*. It can be seenthat, in either case, the absolute value of the blade angle difference|Δβ| in the distribution pattern is less than 10 degrees, indicatingthat the blade angle distribution patterns at the hub surface and at thecasing surface of a blade are substantially similar along any blade.

However, actual flow fields in the diffuser section in an operating pumpare composed of complex three-dimensional flow patterns, and thefrictional effects along the walls on the flow passage producelow-energy fluids which tend to accumulate at the corner regions of thesuction surface and the hub surface due to the secondary flows action.In the conventional designs, a smooth merging of flow passage isproduced by choosing the blade angle distribution as described above.However, because the three-dimensional flow fields are not taken intoconsideration, it has been difficult to prevent a large-scale flowseparation from being generated at the corner or blade root regionswhere the hub surface meets with the suction surface of the blade.

FIG. 16 is a schematic plan view of secondary flows generated on thesuction surface of the blade, while FIG. 17 is a schematic plan view ofthe secondary flow patterns generated on the hub surface in theconventional technology. The low-energy fluids accumulated at the bladeroot regions of the diffuser section do not have sufficient kineticenergy to overcome the pressure rise in the diffuser section, and as aresult, flow separation and reverse flow occur in these blade rootregions as illustrated in FIG. 17.

In the following, the problems encountered in the conventional diffusersection designs will be explained in further detail with reference to athree-dimensional viscous flow analysis. FIG. 18A shows contour lines ofthe static pressure distribution diagram on the suction surface of theblade, and FIG. 18B shows the contour lines of the total pressuredistribution diagram in the flow passage section at a non-dimensionaldistance m*=0.59, and FIGS. 19A and 19B show the predicted velocityvectors close to the suction surface and the hub surface.

As shown in FIG. 18A, in the conventional diffuser section, the contourlines in the entry section of the suction surface (region A) are roughlyparallel to the flow passage P. The flow streams having lost theirkinetic-energy through the frictional effects along the blade wall arenot able to resist the adverse pressure gradient, and generatessecondary flows along the contour lines in the static pressuredistribution diagram, as shown in FIG. 19A.

Because the flow velocity is high in the diffuser entry section,especially near the suction surface, a large friction loss is generatedon the blade walls, and the low-energy fluids are drawn by the secondaryflows on the suction surface and accumulate in the corner regions(region B) formed between the downstream hub section and the suctionsurface.

As can be understood from the dense distribution of the contour linesshown in FIG. 18A, the adverse pressure gradient is high at the cornerregion B, thus generating a large-scale flow separation as illustratedin FIGS. 19A and B, thereby causing a significant loss in the pumpingefficiency. This situation becomes more acute, especially when the pumpis made compact, because the loading on the blade increases and leads toan increase in the adverse pressure gradient, so the pump becomes evenmore sensitive to the separation phenomenon. These are some of the basicreasons that have prevented the conventional technology from makingcompact and high efficiency pumps.

SUMMARY OF THE INVENTION

It is an object of the present invention to provide a highly efficientmixed flow pump by optimizing secondary flows in the diffuser section soas to prevent flow separation which is likely to occur in the cornerregion of the flow passage of the diffuser section.

The object has been achieved in a mixed flow pump comprising a casinghaving an axis and defining an impeller section and a diffuser sectiondisposed downstream of the impeller section. The impeller sectioncomprises an impeller rotating about the axis. The diffuser section hasa hub and stationary diffuser blades, wherein the diffuser blades areformed so that an angular difference, between a hub blade angle and acasing blade angle, is chosen to conform to a specific distributionpattern along a flow passage of the diffuser section. Accordingly, bychoosing an appropriate design of the blade angle of the diffuserblades, a suitable pressure distribution pattern along the flow passagein the diffuser section is obtained by optimizing secondary flows.

In the mixed flow pump presented, the blade angle may be defined interms of an angle between a circumferential tangent line at a point onthe blade surface at a level of hub surface or casing surface and atangent line of a center line of a cross section of the blade along thehub surface or casing surface, and the specific distribution pattern issuch that a hub blade angle is greater than a casing blade angle in awide range of the flow passage. Accordingly, the pressure rise along thehub surface is completed before the pressure rise along the casingsurface so that the flow speed reduction along the hub surface iscompleted before the flow speed reduction on the casing side, therebyenabling the static pressure recovery on the hub side to supercede therecovery on the casing side of the pump.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a perspective drawing of essential parts of an embodiment of amixed flow pump of the present invention;

FIG. 2 is a graph showing a blade angle distribution pattern in adiffuser section of the pump of the present invention;

FIG. 3 is a graph showing a comparison of the differences in the bladeangles along a flow passage in the pump according to an embodiment ofthe present invention and the conventional pump;

FIG. 4A shows contour lines of pressure distribution on a suctionsurface of the blade in the flow passage in the diffuser section in thepump according to an embodiment of the present invention;

FIG. 4B shows contour lines of a total pressure distribution diagram ina circumferential cross section of the flow passage section at anon-dimensional distance m*=0.59 in the diffuser section in the pumpaccording to an embodiment of the present invention;

FIGS. 5A and 5B are velocity vectors of flow fields in the diffusersection in the pump according to an embodiment of the present invention;

FIG. 6A shows contour lines of pressure distribution in a mixed flowpump of the conventional design;

FIG. 6B shows contour lines of the pressure distribution in a mixed flowpump of the present invention;

FIGS. 7A and 7B are graphs to show the performance of the mixed flowpump of the present invention in comparison with the conventional one;

FIGS. 8A˜8F are graphs showing the differences in the diffuser bladeangles along the flow passage of the present invention from the entry toexit sections at different specific speeds;

FIG. 9A is a graph showing distribution of blade angle difference Δβbefore amendment for the mixed flow pumps of the present invention;

FIG. 9B is a graph showing distribution of blade angle difference Δβ*after amendment for the mixed flow pumps of the present invention;

FIG. 10 is a graph showing the relationship between the specific speedsand the non-dimensional distance of the location of the maximum bladeangle difference for the mixed flow pumps shown in FIGS. 8A˜8F;

FIG. 11 is a graph showing the maximum blade angle difference as afunction of the specific speed for the mixed flow pumps shown in FIGS.8A˜8F;

FIG. 12 is a schematic cross sectional view of a conventional mixed flowpump;

FIG. 13A is a drawing to illustrate the definition of the blade angle βon a casing surface of the diffuser blade;

FIG. 13B is a drawing to illustrate definition of the coordination on ameridional surface of the diffuser blade;

FIG. 13C is a drawing to illustrate the coordination and the blade angleβ on an axisymmetrical surface of the diffuser blade section;

FIG. 13D is a drawing to illustrate the definition of the amended bladeangle β* of the diffuser blade when it is slanted;

FIG. 14A is a graph showing a distribution pattern of blade angles inthe diffuser section of a conventional mixed flow pump;

FIG. 14B is a graph showing a distribution pattern of average bladeangles in the diffuser section of the mixed flow pump of the presentinvention compared with a conventional one;

FIG. 15 is a graph showing the blade angle difference Δβ as a functionof the non-meridional distance m* in the conventional mixed flow pump;

FIGS. 16 is an illustration of the secondary flow patterns on thesuction surfaces of the diffuser blade in the conventional mixed flowpump;

FIG. 17 is a plan view of the secondary flow patterns on the hub surfaceof the diffuser section in the conventional mixed flow pump;

FIG. 18A shows the contour lines of the pressure distribution on thesuction surface of the blade in the flow passage in the diffuser sectionin the conventional mixed flow pump;

FIG. 18B shows the contour lines of the total pressure distributiondiagram in a circumferential cross section of the flow passage sectionat a non-dimensional distance m*=0.59 in the diffuser section in theconventional mixed flow pump; and

FIGS. 19A and 19B show velocity vector patterns in the diffuser sectionof the conventional mixed flow pump.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

FIG. 1 shows essential components of a mixed flow pump of an embodimentaccording to the present invention. The essential feature of theinvention resides in a configuration of diffuser blades 20 in a diffusersection 14. The blade angles of the blades 20 of the pump aredistributed along the meridional surfaces as shown in FIG. 2 in whichthe horizontal axis relates to the non-dimensional distances along theflow passage, and the vertical axis relates to the blade angle β asdefined in FIG. 13A. As can be understood from this, the blade angleβ_(h) of the blade 20 on the hub surface increases gently to a vicinityof a point given by a non-dimensional distance m*=0.5, but thereafter itincreases rather sharply. On the other hand, the blade angle β_(c) onthe casing surface increases gently at about the same rate as β_(h) to anon-dimensional distance m*=0.4 and continues to increase at about asimilar rate to a non-dimensional distance m*=0.75, and thereafterincreases quite sharply.

The result is that, as shown in a comparative diagram in FIG. 3, theblade angle difference Δβ between the hub blade angle β_(h) and thecasing blade angle β_(c) is about the same in the front half of thediffuser flow passage P, but in the rear half of the diffuser flowpassage P, the hub blade angle β_(h) is larger than the casing bladeangle β_(c). In this example, the blade angle difference Δβ increasesrapidly from a point at m*=0.5, and the difference reaches a peak valueof about 30 degrees at m*=0.75. It can be recognized that this angulardistribution pattern is significantly different from the conventionaldistribution pattern shown in FIG. 15.

FIGS. 4A, 4B and 5A, 5B show predicted pressure distribution patternsand velocity vectors in the flow passage P in the diffuser section 14 ofthe present mixed flow pump, computed by using a three-dimensionalviscous flow analysis. The contour lines of the static pressures in theentry section (region A′) shown in FIG. 4A are formed aboutperpendicular to the passage P, and the secondary flows flowing alongthe contour lines flow towards the hub surface as shown in FIG. 5A.Therefore, due to the changes in the secondary flow pattern, thehigh-loss fluid which would have been accumulated in the corner regionof the diffuser section in the conventionally designed diffuser ispassed over the corner region and is accumulated in a region D′ on thehub side in the mid-pitch location of the flow passage. The high-energyfluid flowing in the casing-side flows into the corner region (regionC′, refer to FIG. 4B), and because the adverse pressure gradient in thisregion is small (region B′, refer to FIG. 4A), the flow separationgenerated on the, hub surface is shrunk, as can be confirmed in FIG. 5B,thereby improving the flow fields significantly.

In the present distribution pattern of the blade angles, the increasesin the blade angle β_(h) on the hub surface precedes that on the casingsurface. The result is that the pressure increase on the hub-side iscompleted before the pressure increase is completed on the casing-side.Accordingly, the present diffuser enables the establishment of staticpressure contour lines which are nearly perpendicular to the flowpassage P as illustrated in a comparative flow pattern shown in FIG. 6B,compared with a conventional flow pattern shown in FIG. 6A. Furthermore,because the pressure increase is completed in the front half of theblade where the boundary layer thickness is small and the resistance toflow separation is high, the present flow fields enable the moderationof the adverse pressure gradient in the region B′ where the boundarylayer thickness is large and the resistance to flow separation is low,thereby realizing a suppression effect of the flow separationphenomenon.

FIGS. 7A and 7B show a performance comparison of a mixed flow pump withthe present blade design with an equivalent mixed flow pump with theconventional blade design with a specific speed 280 (m, m³/min, rpm). Itcan be seen that the present design of the blade angle distribution hasproduced significant performance improvements over the blade angledistribution used in the conventional design. The specific speed Ns isgiven by the following equation:

Ns=NQ ^(0.5) /H ^(0.75)  (2)

where N is a rotational speed of the impeller in rpm, Q is a design flowrate in m³/min and H is the total head of the pump in meters at thedesign flow rate.

FIGS. 8A˜8F show examples of the present design diffuser at specificspeeds ranging from 280 to 1,000 (m, m³/min, rpm). Each drawing showsthree or four distribution curves of the blade angle difference Δβ ofthe diffuser blades 20 having different meridional surface shapes.Although differences in the maximum blade angles caused by thedifferences in the meridional surface shapes can be observed, thecharacterizing feature of the present diffuser design, that generallythe blade angle difference increases sharply along the flow passage,from the entry side to the exit side of the diffuser section, is clearlyvisible in each example.

It can be seen that the peak point, where the blade angle difference Δβis a maximum, shifts from the rear half of the flow passage to the fronthalf of the flow passage, as the specific speed increases. It will alsobe noted that the maximum blade angle difference decreases at higherspecific speeds. Also, the rise point, where the blade angle differencebegins to increase, is where non-dimensional distance m*=0.4 at aspecific speed of 280 while at the specific speeds of over 400, theblade angle difference begins to increase near the leading edge of thediffuser section. As the specific speed decreases, the load on thediffuser blades increases, therefore, in order to prevent the flowseparation phenomenon at low specific speeds, it is necessary that alarger blade angle difference Δβ is realized. At all specific speeds,after the blade angle difference reaches a maximum, the differencediminishes quickly towards the trailing edge where non-dimensionaldistance m* is 1, and at the trailing edge of the diffuser section 14,the difference is almost zero.

The circumferential coordinates θ_(TE) at the trailing edge location ofthe diffuser section are often made to be identical, from the viewpointof ease in manufacturing, on the hub (θ_(TE)=θ_(TE,h)), and on thecasing (θ_(TE)=θ_(TE,c)), so that the trailing edges are oriented in theradial direction. If the blades at the trailing edges are slanted in thecircumferential direction (i.e., θ_(h)≠θ_(c)), performance improvementscan be obtained if the distribution of the blade angle difference isamended into an equivalent one satisfying θ_(h) =θ_(c) condition. Suchamendment is conducted according to the following equations:

θ*_(h)=θ_(h) +m*·Δβθ _(TE)  (3)

tan β*_(h) =dm/d(rθ*_(h))  (4)

Δβ*=β*_(h)−β_(c)  (5)

where θ_(h) is a circumferential coordinate of the center line on thehub surface of a blade; Δθ_(TE) is the difference in the circumferentialangles at the trailing edge between the hub and the casing(θ_(TE,c)−θ_(TE,h)); θ*_(h) is circumferential coordinate of the centerline of the hub surface after the amendment; β*_(h) is the blade angleon the hub surface after the amendment; and Δβ* is the blade angledifference after the amendment (refer to FIG. 13D).

FIGS. 9A and 9B show the effects of varying the blade slant angleΔθ_(TE) from about −6 to 17 degrees in an embodiment of a mixed flowpump with a specific speed of 400 (m, mn³/min, rpm). The distribution ofthe blade angle differenceΔβ before the amendment is different indifferent blade slant angles Δθ_(TE) as shown in FIG. 9A, but after theamendment process according to the above equations, the distribution ofthe blade angle difference Δβ* becomes substantially the same, therebyconfirming the fact that the amendment process for Δβ* is universallyapplicable. It should be clear from Equation (1), when θ_(h)=θ_(c),i.e., Δθ_(TE)=0, then Δβ*=Δβ.

FIG. 10 summarizes non-dimensional distance, designated as m*_(p), wherethe blade angle difference Δβ* shows a maximum value in various examplesas a function of the specific speeds, and FIG. 11 summarizes the maximumvalues of the blade angle difference Δβ*. In the figures, the solidcircles  refer to the cases of slanted blades (θ_(h)≠θ) at the trailingedges of the diffuser section.

As shown by the solid lines in the figures, the lower limit m*_(p,min)and the upper limit m*_(p,max) for the non-dimensional distancemaximizing the values of the blade angle difference Δβ*; and the lowerlimit Δβ*_(min) and the upper limit Δβ*_(max) for the maximum bladeangle difference; are given by the following equations:

m* _(p,min)=0.683−0.0333·(Ns/100)  (6)

m* _(p,max)=1.12−0.0666·(Ns/100)  (7)

Δβ*_(min)=30.0−2.50·(Ns/100)  (8)

Δβ*_(max)=53.3−3.33·(Ns/100)  (9)

FIG. 14B shows an example of a pump with a specific speed of 280 (m,m³/min, rpm), and compares the distribution patterns of the averageblade angles at mid-span location in the present diffuser section (referto FIG. 2) and those in the conventional diffuser section (refer to FIG.14A, case N). Clearly demonstrated, although the two cases share roughlysimilar distribution patterns of the average blade angles, theconventional pump shows a large degree of flow separation as shown inFIGS. 19A and 19B, whereas the present pump shows suppression of flowseparation as shown in FIGS. 5A and 5B, and the pump performance issignificantly improved as shown in FIGS. 7A and 7B. These resultsdemonstrate convincingly that what is important is not the average bladeangle distribution pattern but it is the difference in the blade angleon the hub and casing that determines the pump performance. It can beunderstood that a major cause of degradation in the pump performance isthat the conventional diffusers have placed emphasis on a smoothtransition of the blade angle distribution pattern from the entry to theexit, and no special consideration has been given to the important roleof the changes in the blade angle difference distribution patternbetween the hub surface and the casing surface of the blades from theentry to the exit of the diffuser section, as in the present invention.

In brief summary, the present invention has demonstrated that anefficient mixed flow pump can be produced by designing the diffuserblade so that the difference in the blade angle, at the hub and at thecasing, changes according to a specific distribution pattern, along theflow passage from the entry-side to the exit-side in the diffusersection. The distribution pattern is determined by the criteria tooptimize the generation of secondary flows and to prevent separation atthe corners of the flow passage cross section in the diffuser section.

What is claimed is:
 1. A mixed flow pump comprising a casing having anaxis and defining an impeller section and a diffuser section disposeddownstream of said impeller section, said impeller section comprising animpeller adapted to rotate about the axis, said diffuser section havinga hub and a plurality of stationary diffuser blades, wherein saiddiffuser blades are formed so that an angular difference between a hubblade angle and a casing blade angle conforms to a specific distributionpattern along a flow passage of said diffuser section, the specificdistribution pattern being such that the angular difference increasesalong the flow passage, reaches a maximum difference, and then decreasesalong the flow passage.
 2. A mixed flow pump according to claim 1,wherein the hub blade angle and the casing blade angle are each definedin terms of an angle between a circumferential tangent line at a pointon a surface of one of said diffuser blades at a level of a hub surfaceor a casing surface, respectively, and a tangent line of a center lineof a cross section of one of said diffuser blades along said hub surfaceor said casing surface, respectively, and the specific distributionpattern is such that an increase in the hub blade angle on said hubsurface precedes that of the casing blade angle on said casing surfacealong the flow passage.
 3. A mixed flow pump according to claim 2,wherein a lower limit of a maximum value in a distribution pattern ofamended blade angle differences (Δβ*), defined by a difference(β*_(h)−β_(c)) between the hub blade angle amended to be an amended hubblade angle (β*_(h)) on a hub end of one of said diffuser blades and thecasing blade angle (β_(c)) on a casing end of one of said diffuserblades, is located on an exit-side of a location with a non-dimensionaldistance m*_(p,min) represented by an equation:m*_(p,min)=0.683−0.0333·(Ns/100).
 4. A mixed flow pump according toclaim 3, wherein an upper limit of the maximum value in the distributionpattern of the amended blade angle differences (Δβ*) is located on anentry-side of a location with a non-dimensional distance m*_(p,max)represented by an equation: m*_(p,max)=1.12−0.0666·(Ns/100).
 5. A mixedflow pump according to claim 2, wherein a lower limit of a maximum valuein a distribution pattern of amended blade angle differences (Δβ*),defined by a difference (β*_(h)−β_(c)) between the hub blade angleamended to be an amended hub blade angle (β*_(h)) on a hub end of one ofsaid diffuser blades and the casing blade angle (β_(c)) on a casing endof one of said diffuser blades, is not less than a value given by anequation: Δβ*_(min)=30.0−2.50·(Ns/100).
 6. A mixed flow pump accordingto claim 5, wherein an upper limit of the maximum value in thedistribution pattern of the amended blade angle differences Δβ* is notmore than a value given by an equation: Δβ*_(max)=53.3−3.33·(Ns/100). 7.A mixed flow pump according to claim 1, wherein a lower limit of amaximum value in a distribution pattern of amended blade angledifferences (Δβ*), defined by a difference (β*_(h)−β_(c)) between thehub blade angle amended to be an amended hub blade angle (β*_(h)) on ahub end of one of said diffuser blades and the casing blade angle(β_(c)on a casing end of one of said diffuser blades, is located on anexit-side of a location with a non-dimensional distance m*_(p,min)represented by an equation: m*_(p,min)=0.683−0.0333·(Ns/100).
 8. A mixedflow pump according to claim 7, wherein an upper limit of the maximumvalue in the distribution pattern of the amended blade angle differences(Δβ*) is located on an entry-side of a location with a non-dimensionaldistance m*_(p,max) represented by an equation:m*_(p,max)=1.12−0.0666·(Ns/100).
 9. A mixed flow pump according to claim1, wherein a lower limit of a maximum value in a distribution pattern ofamended blade angle differences (Δβ*), defined by a difference(β*_(h)−β_(c)) between the hub blade angle amended to be an amended hubblade angle (β*_(h)) on a hub end of one of said diffuser blades and thecasing blade angle (β_(c)) on a casing end of one of said diffuserblades, is not less than a value given by an equation:Δβ*_(min)=30.0−2.50·(Ns/100).
 10. A mixed flow pump according to claim9, wherein an upper limit of the maximum value in the distributionpattern of the amended blade angle differences Δβ* is not more than avalue given by an equation: Δβ*_(max)=53.3−3.33·(Ns/100).